Hydraulic ball pumps and motors



June 8, 1965 D. FIRTH ETAL 3,187,581

I HYDRAULIC BALL PUMPS AND MOTORS Filed Jah. 12 1962 5 Sheets-Sheet 1June: 8', 11965 D. FIRTH ETAL HYDRAULIC BALL PUMPS AND MOTORS 5Sheets-Sheet 2 FilediJEm. 12, 1962' June 1965 D. FIRTH ETAL 3,187,681

HYDRAULIC BALL PUMPS KN-D MOTORS Filed Jan. 12, 1962 5 Sheets-Sheet 5June :8, 19165 D. FIRTH ETAL 3,137,681 I HYDRAULIC BALL PUMPS AND MOTORS5 Sheets-Sheet 4 Filed Jan. .32., v1.962

June 8, 1965 FlRTH ETAL 3,187,681

HYDRAULIC BALL PUMPS AND MOTORS Filed Jan. 12, 1962 5 Sheets-Shea 5United States Patent 0 3,187,681 RAULIC BALL PUMPS AND MQTGRS DonaldFirth and Roger Harvey Yorke Hancock, East Kilhride, Glasgow, Scotland,assignors to National Research Development Corporation, London, England,a British corporation Filed Jan. 12, 1962, Ser. No. 165,9134 Claimspriority, application Great Britain, Han. 27, 1961, 3,2ltl/61 4 Claims.(Ci. 1433-161) The present application is a continuation-in-part of ourapplication Serial No. 39,538, filed June 29, 1960 now abandoned.

This invention relates to hydraulic ball pumps and motors in which arotary cylinder block has one or more (usually about six) cylindersopening through its periphery, the piston of each cylinder beingconstituted by a ball which runs on an eccentric track or cam ringsurrounding the cylinder block. is divided into two halves by a flatdiametral section of a fluid-tight pintle, each cylinder being open atits inner end to the cavity. One-half of the cavity is a fluid inletduct and the other half a fluid outlet or delivery duct. During theoutward travel of a ball, its cylinder is open to the inlet duct andduring its inward travel the cylinder is open to the outlet duct.

An advantage of the ball pump over a piston type pump is its simplicityand hence its cheapness. It would be a great advantage to be able to usea ball pump in a high pressure system, since it would considerablyreduce the cost of such a system. Up to the present time, however, ithas not been possible to use a ball pump to deliver oil at highpressure, e.g. over about 300 lb./sq. in. It has been found that whenthe speed of rotation and the delivery pressure of a ball pump areincreased, the pump normally seizes and becomes damaged to such anextent that it is not possible to determine the cause of the seizure.Theories have been advanced to explain why a ball pump could not operateat high speeds and pressures. For example, it has been thought thatsince the bearing surface between a ball and its cylinder is in the formof a sliding line contact, it would be impossible to lubricateadequately the bearing surface at high speeds.

The oil which leaks past a ball between its circumference and thecylinder Wall, serves as a lubricant both between the ball and thecylinder Wall and also between the ball and the cam ring. The leakageincreases with increased pump pressure, and the ratio of leakage oil tooil pumped can be kept to a minimum if for any given design requirementthe balls are kept as large as possible and the speed of rotation at amaximum. Nevertheless, seizure limits the speed of rotation.

Careful investigation has now shown the reasons for the seizure ofconventional ball pumps at high outputs. Chief among these reasons isthe breakdown of the lubricant film at two diametrically opposite pointson the surface of the ball. These points are the poles about which theball rotates by reason of its contact with the cam ring. Oil is thrownoff the surface at these points by centrifugal force at higher speeds ofrotation, and dry contact is established between the ball and thecylinder wall. This has two effects. The first is that the ball wearsits own surface and that of the cylinder wall, so that the initialsmoothness breaks down locally. The second is that the roughening of themating surfaces between the ball and the cylinder sets up highfrictional drag on the ball which then begins to slide over theeccentric track of the cam ring. Since the latter has hitherto been flatin many practical constructions of pump or motor, there is a high localloading on the ball and the track, and breakdown An axial cavity in theblock "ice 7 of the contacting surface rapidly builds up as soon assliding begins.

These deleterious effects are cumulative and are equally applicable topumps and motors. They quickly lead'to seizure of the machine.

The present invention is a ball pump or motor in which theabove-mentioned defects are materially reduced or eliminated.Accordingly, an improved radial hydraulic machine preferably has atleast one of the contacting surfaces between a ball and its cylindercoated with an anti-friction wear-resistant hard material having acoefficient of friction not greater than 0.2.

Advantageously, the anti-friction material is tungsten carbide, and thecylinder wall is conveniently constituted by a liner shrunk into thecylinder, both the initial cylinder bore and the external surface of theliner being complementarily stepped to present a slightly largerdiameter over the inner end section of the cylinder and liner. Thestepped portions of the liner and cylinder bore may be relativelyproportioned so that the liner is positively locked in the cylinder byaxial pressure at the shoulder.

It has however been discovered that longitudinal clamping of the lineris not essential and hence the accuracy to which the length of theenlarged diameter sections of the cylinder and liner have to be machinedin order to establish axial compression can be materially relaxed, withconsequent advantages in cost and time of machining. Therefore in apreferred form the invention comprises the combination of a cylinderbore and a liner both of which are stepped to present enlarged diametersections at the inner end of the cylinder, the radial height of the stepbetween the sections on the liner being such that the enlarged diametercan be passed through the smaller diameter section of the cylinder borewhen the cylinder is at a temperature sufiiciently higher than that ofthe liner, the axial length of the enlarged diameter section of thecylinder being greater than the axial length of the correspondingsection of the liner. Preferably, a small radial clearance is alsoprovided between the larger diameter section of the cylinder bore.Conveniently, the diameters of the smaller diameter sections of thecylinder bore and liner are chosen so that the liner is a smallinterference fit in the cylinder bore over this section sufiicient toprovide a reasonable radial grip and to ensure that the shoulder on theoutside of the liner can engage the corresponding shoulder in thecylinder bore to prevent unintentional withdrawal of the liner from thebore under operating conditions.

Preferably, the thrust between a ball and its cylinder on a pumpingorpower stroke is reduced by skewing or inclining the axis of the cylinderto the true radius of the cylinder block through the centre of the ballso that a greater part of the thrust exerted by the cam ring on the ballis absorbed by the oil under pressure in the cylinder.

Practical embodiments of the invention will now be particularlydescribed, by way of illustration only, with reference to theaccompanying drawings wherein:

FIGURE 1 is a longitudinal section through a four-ring, eight-cylinderpump or motor, some parts being omitted for purposes of illustration,

FIGURE 2 is a section, partly in elevation, on the line IIII of FIGURE1;

7 FIGURE 3 is a section through a cylinder modified in accordance withthe present invention; 1

FIGURE 4 is a schematic sectional perspective View of a cylinder and itsball illustrating certain working conditions;

FIGURES 5 and 5A are vector diagrams of the forces acting on a ball in atypical machine having a conventional cylinder arrangement;

FIGURES 6 and 6A are similar diagrams relating to a modified cylinderarrangement according to the present invention;

FIGURE 7 is a schematic sectional view of a detail;

FIGURE 8 is a fragmentary cross section in generally schematic formthrough a modified cylinder and liner, and

FIGURE 9 is a fragmentary sectional view similar to FIGURE 8illustrating a further modification.

Referring first to FIGURES l and 2, the ball pump or motor showncomprises a rotary cylinder block 1 having four rings of nine generallyradial cylinders-2 within each of which works a ball 3. The halls run oneccentric ring cams 4, one for each ring of cylinders 2, and each cam,

4 is pivoted on a common spindle 5 in the casing 6 of the machine. Thestroke of each ball 3 in it's cylinder 2 during a complete revolution iscontrollable by adjustment of the eccentricity of its respective cam 4relatively to the axis of the cylinder block 1. This adjustment iseffected by means of a shaft 7 which carries four circular eccentricsheaves 8, one for each cam 4. Each sheave 8 lies snugly between a pairof opposed horns 9 bolted to the respective cam 4, and theeccentricities of the sheaves are alternately arranged at 180 withrespect to each other. The throws of the sheaves are equal, so that asthe shaft 7 is rotated, adjacent cams 4 are adjusted equally in oppositesenses.

The shaft 7 is driven by a pinion 10 from a rack 11 which may bemanually or automatically controlled. Adjustment of the stroke of eachball piston 3 can thus be made dependent on the load demand on the othermachine. Since thenumber of pistons on load at any one instant variescyclically during each revolution, the cams 4 are clamped by hydraulicplungers 12 (FIG. 2) working in cylinders 13 set radially in the casing6 at approximately 90 to the horns 9. a

The inlet and outlet port events of the machine are controlled in theconventional manner by a pintle 14 having a pair of longitudinal ducts15, 16. Each duct is connected to the upper or lower of a pair oftransverse grooves 17, 18 in the pintle, each pair of grooves being inregister with the ports 19 of a respective ring of cylinders 2. Thelongitudinal duct communicates, through ports 20, 21, alternately withthe upper and lower grooves 17, 18 whilst the lower duct 16communicates, in the opposite alternation, with these grooves throughports 22, 23.

In order to promote smooth running of the machine, the instant ofclosing and opening of each port 19 can be varied by the angularadjustment of the pintle 14. This is effected by means of an eccentricpin 24 carried in a.

stud 25 rotatably mounted in one end of the casing 6.

and having a squared head 26 for engagement by a spanner or otherdevice. The eccent'ricpin 24 Works in a longitudinal groove 27 in thesurface of a flange 28 on the projecting end of the pintle 14, and thestud 25 is both located and locked by a grub screw 28.

As shown in both FIGURES 1 and 2, each cylinder 2 is sleeved by means ofa hardened and accurately ground liner 30. This liner is shown in FIGS.1 and 2 as a simple cylinder for convenience of drawing on a relativelysmall scale, .but this is not the true form of the liner. In actualpractice, both the cylinder bore 29 (FIG. 3) and the external surface ofthe liner 30 are complementarily stepped or shouldered at31,approximately' mid-way along the length of the bore. This step orshoulder is shown exaggerated in FIG.-3 for clarity of illustrationonly. of the liner 30 has an external diameter such that, after theliner 30 has been cooled in a refrigerator while the block 1 has beenheated to a relatively high temperature, the enlarged end 32 can justpass through the outer end of the bore 29.

The axial height of the enlarged end 32 of the sleeve and of thecorresponding undercut in the cylinder bore 29 between the shoulder 31and the inner end are so propor- In practice, the enlarged end 32.

tioned that, on cooling of the block 1 after insertion of a liner 30,the enlarged end '32 is subjected to axial compression to ensure thatthe liner is securely locked in position. This arrangement has theadvantage of avoiding the necessity for relying on frictional gripbetween the liner 30 and the cylinder bore 29, which normally calls forrelatively high radial compressive loads. These in turn lead todistortion of the bore, which must then be reground in situ. In theaforesaid shouldered construction according to the present invention,radial compression is not relied on to provide frictional locking of theliner, and both the cylinder bore 29 and the liner 30 are thereforeinitially machined to greater tolerances on diameter so as to permitsignificant reductions in the radial compressive forces set up onshrinkage as the parts cool.

FIGURE 4 illustrates the operating conditions of aball 3 in its cylinder2. Superimposed on a linear reciprocation along the axis of the cylinder2 is a'rotation about a perpendicular axis XX. This rotation isindicated by the arrows s. The chain line Q represents the equatorialline of contact between the ball 3 and cylinder 2, and this lineconstitutes the sealing surface which holds back the oil' under pressureduring the compression stroke. Where the line Q intersects the polaraxis XX are the two poles P of the ball which merely slide up and downthe cylinder wall during each stroke. They are the only points on theball which only remain in contact with the cylinder wall.

As the speed of rotation of the ball 3 increases, oil leaking across theline of contact Q tends to be thrown off a zone around each of the polesP. I-Ience dry contact becomes established at these points between theball and the cylinder wall during each stroke of the ball, andfrictional drag and wear commence at relatively low speeds of rotationunless at least one of the contacting surfaces is itself of ananti-friction material. Experiment has shown that, provided thecoefiicient of friction at the cylinder wall is kept below 0.2andpreferably below 0.18--a substantial reduction in both these tendenciesis exhibited. Such low figures of friction are obtainable by the use oftungsten carbide for the linear 30-or at least its working surface-andoptionally also for the surface of the ball 3. Either component can beof steel flame-plated with tungsten carbide. The tungsten carbidesurface is ground and lapped to size and a liner 30 made .in this waypresents a verylow-friction surface to a hardened steel ball 3 enablingit to run at higher speeds than heretofore in spite of the onset of drycontact at the poles P.

FIGURES 5 and 5A illustrate diagrammatically the forces acting on a ball3 during the pumping stroke, i.e. at an intermediatepoint between bottomand top dead centres. Due to the eccentricity of the track on the camring 4, the total reaction force U (neglecting friction) exerted on theball 3 is not radial to the cylinder block 1 except at bottom and topdead centres, but acts on the line BC through the centre B of the balland the centre C of the cam 4. The cam eccentricity is the distance ebetween the cam centre C and the centre A of the cylinder block. Thepressure of the hydraulic fluid in the cylinder 2 exerts a force Facting along the axis of the cylinder. The cylinder block thus exerts athrust T on the ball which, neglecting friction at the ball, is thevector sum of the forces F and U. This force T is shown in FIGURE 5A (inwhich the axis of the cylinder 2 is radial) and it can be regarded asacting at the centre of the ball 3, at right angles to the polaraxisX-X. It is the force tending tocause breakdown of the oil film onthe line of contact Q (FIG. 4) between the hall 3 and the cylinder 2other than at the poles P, and since it acts at a distance AB from thecentre A of the cylinder block 1 it produces a torque which must beovercome by the input drive. It. also produces a second frictional dragbetween ball and cylinder which is also materially 3 reduced by thelow-friction tungsten carbide wearing surface.

It is, therefore, desirable to reduce the magnitude of V the force T,and FIGS. 6 and 6A show diagramatically how this can be done. In FIG. 6,B is the centre of the ball 3 and A is the centre of the cylinderblock 1. BC represents the line of the reaction force U, C being thecentre of the cam ring 4. The axis of the cylinder 2 is, however, at anangle to the radius AB, lying along the line BD which thus becomes thenew line of action of the force P which is due to the pressure of theoil being pumped. The component T is accordingly reduced, as shown inFIG. 6A. The torque to be overcome by the input drive is now Pw-l-TUZD),where w is the perpendicular distance from the centre A of the cylinderblock 1 to the cylinder axis BD. The optimum value of 0 is that forwhich where e is the concentricity of the cam ring.

The skewing of the axis BD of the cylinder 2 materially reduces thethrust between the cylinder wall and the ball, so that breakdown of theoil film on the line of contact Q (FIG. 4) away from the poles P isdelayed and permits higher operating speeds.

The rate of wear can be materially retarded if the Hertzian stresses setup in the ball by the fluid pressure during each pumping or power strokeare reduced. This can be achieved by arranging that the track on the camring 4 along which the balls 3 run is grooved at 34 to a radius Rslightly greater than that (R of a ball so that, as the pressure of theoil increases, the local loading on the ball where it makes contact withthe track is proportionately less. A suitable value for R is between1.02 R and 1.07 R

The pintle 14 is angularly adjusted, as described above, so as to delaythe instant of cut-oif of the ports 19 at the ends of the delivery(pump) or power (motor) strokes of their respective pistons 3 so as toallow the pistons to commence their suction or exhaust strokes beforethe ports are opened to the oil at inlet or exhaust pressure, as thecase may be, in the appropriate groove 17 or 13. The extent of thisdelay is chosen so that the residual oil at high pressure in eachcylinder 2 is allowed to expand down to approximately the initialpressure on the next stroke before the port 19 opens, and thus hammeringor shock Waves in the system due to the sudden exposure of the highpressure residual oil to a low pressure chamber can be minimised orobviated. In this way, stresses in the balls 3 are further reduced, anddangerous wear is materially postponed. The efiects of this reduction ofhammering and of the reduction of Hertzian stresses by the provision ofa grooved cam track 34, as already described, are cumulative and, whentaken together, result in a much improved rate of wear.

The control of the angular adjustment of the pintle 7 may be manual orautomatic-the latter depending on the instantaneous output of the pumpor motor.

Refering to FIGURE 8 of the drawings, a cylinder block 41 of, say a ballpump is provided with a number (only one is shown) of radial orsubstantially radial cylinders 42 each having a bore which is stepped atan internal shoulder 43 so as to present an inner end 44 of a largerdiameter than the outer end 45. Into this cylinder is inserted a sleeveor liner 45 whose inner periphery is of constant diameter and whoseouter wall is similarly stepped at 47 so as to present an inner end wallsection 43 or" larger diameter than the outer end wall section 49. Bothare slight interference fits in the respective inner and outer sections44 and 45 of the cylinder bore. In this way the liner is secured againstaccidental removal from the cylinder bore without the necessity forexerting high radial compressive stress on the liner 4-6.

One of the drawbacks of the above arrangement is the necessity for theaccurate internal machining of two nearly equal diameters on thecylinder wall. 7

FIGURE 9 shows a modification in which he larger diameter inner end 44aof the cylinder bore is both longer and of greater diameter than thecorresponding inner end section 48a of the sleeve 460 so that there isclearance both onthe diameter and on the length between the respectiveinner end sections 44a, 48a. At the same time, the outer end section 49of the liner is a small interference lit in the outer end section 45 ofthe cylinder bore and this interference is chosen primarily to retainthe sleeve against radial rather than against axial displacement undermost normal working conditions of the machine. With this arrangement ithas been found easier to hold the parallelism of the bore of the sleeve46. The height of the shoulder 47 on the sleeve 46 is still sufficientto ensure that it can engage the shoulder 43 in the cylinder bore undernormal conditions of use so as to positively resist axially outwarddisplacement of the sleeve.

In each case, the larger diameter liner section 48 or 48a is of anexternal diameter such that it can be passed through the outer section45 of the cylinder bore when the parts are at sufiiciently differenttemperatures.

Although satisfactory results have been obtained by the use of atungsten carbide liner 30 in conjunction with a hardened steel ball 3,the latter may, if desired, be of a harder material, or may have itssurface flame-plated with tungsten carbide to reduce the coefficient offriction between ball and cylinder still further. It has also been foundthat the optimum tolerance on the difference of diameter between a ball3 and its cylinder 30 is 0.00015 to 0.0004 inch.

In conventional ball pumps and motors as hitherto constructed, it hasnot been possible to exceed a working pressure of the order of 300lb./in. nor a size of 2 to 3 hp. Working speed at these ratings has beenlimited to less than 1000 r.p.m.usually a maximum of 800 r.p.m. with amaximum working life of considerably less than 1000 hours. By contrast,a prototype machine according to the present invention has been run atspeeds up to 3000 rpm. with a working pressure of 2000 lo./in. at apower of 50 hp. for 6000 hours. The machine was then stripped down andno measurable wear found. It was re-assembled and has been put back intoservice.

We claim:

1. An hydraulic radial machine comprising a casing; a rotary cylinderblock journalled in the casing; a plurality of substantially radialcylinder bores in the cylinder block, each bore having a cylindricalwall containing an annular groove having axially-spaced side walls; aring cam surrounding the cylinder block with radial clearance therefrom;a ball piston located in each cylinder and bearing on the ring cam; apintle located in an axial bore in the cylinder block and controllingthe port events of each cylinder; and a liner with a hard, low-friction,internal surface in each cylinder bore in the block, said liner having acylindrical external surface which is an interference fit in thecylinder bore and an external flange disposed in said groove between theside walls thereof and with an interference fit against the bottomthereof.

2. An hydraulic radial machine comprising a casing; a rotatry cylinderblock journalled in the casing; a plurality of substantially radialcylinder bores in the cylinder block, each bore having a cylindricalwall containing an annular groove having axially-spaced side Walls; aring cam surrounding the cylinder block with radial clearance therefrom;a ball piston located in each'cylinder and hearing on the ring cam; apintle located in an axial bore in the cylinder block and controllingthe port events of each cylinder; and a liner with a hard, low-friction,internal surface in each cylinder bore in the block, said linercomprising a portion with a cylindrical outer surface which is aninterference fit in the cylinder bore and another portion ofsubstantially the same axial length as the firstsaid portion anddisposed in said groove with an interference fit against the bottomthereof.

3. An hydraulic radialmachine comprising a casing; a rotary cylinderblock journalled in the casing; a plurality of substantially radialcylinder bores in the cylinder block, each bore being cylindrical andhaving'adjacent its radially inner end an annular groove withaxiallyspaced side walls; a ring cam surrounding the cylinder block withradial clearance therefrom; a ball piston located in each cylinder andbearing on the ring cam; a pintle located in an axial bore in thecylinder block and controlling the port events of each cylinder; and aliner with a hard, low-friction, internal surface in each cylinder borein the block, said liner having a cylindrical external surface which isan interference fit in the cylinder bore and a portion of largerdiameter at its axially innermost end which is disposed in said groovewith an interference fit against the bottom thereof.

4. An hydraulic radial machine comprising a casing; a rotary cylinderblock journalled in the casing; a plurality of substantially radialcylinder bores in the cylinder block, each bore being cylindrical andhaving adjacent its radially inner end an annular groove withaxiallyspaced side walls; a ring cam surrounding the cylinder block withradial clearance therefrom; a ball piston located in each cylinder andbearing on the ring cam; 21 pintle located in an axial bore in thecylinder block and controlling the port events of each cylinder; and aliner bore in the block, said liner consisting of two portions withcylindrical outer surfaces and of substantially equal length, oneportion having an interference fit against the cylinder bore and theother portionbeing disposed in said groove with an interference fitagainst the bottom thereof.

References Cited by the Examiner UNITED, STATES PATENTS 1,057,741 4/13Keen 123-193 1,325,434 12/19 Carey et al 103-161 1,347,476 7/20 Allyne29-1564 1,788,950 1/31 Heiner 309-3 2,246,392 6/41 Spreng et a1. 309-32,387,971 10/45 Aspin et al. 309-3 2,455,457 12/48 Whittfield 309-32,646,755 7/53 Joy 103-161 2,815,718 12/57 Avery 103-161 2,823,619 2/58May 103-216 2,838,003 6/58 Joy 103-161 2,855,858 10/58 Larsen et al.103-161 2,862,455 12/58 Ferris 103-161 2,952,215 9/60 Deschamps 103-216FOREIGN PATENTS 924,602 3/47 France. 574,991 1/ 46 Great Britain.

. LAURENCE V. EFNER, P i E with a hard, low-friction, internal surfacein each cylinder r mary xammer JOSEPH H. BRANSON, JR., Examiner.

1. AN HYDRAULIC RADIAL MACHINE COMPRISING A CASING; A ROTARY CYLINDERBLOCK JOURNALLED IN THE CASING; A PLURALITY OF SUBSTANTIALLY RADIALCYLINDER BORES IN THE CYLINDER BLOCK, EACH BORE HAVING A CYLINDRICALWALL CONTAINING AN ANNULAR GROOVE HAVING AXIALLY-SPACED SIDE WALLS; ARING CAM SURROUNDING THE CYLINDER BLOCK WITH RADIAL CLEARANCE THEREFROM;A BALL PISTON LOCATED IN EACH CYLINDER AND BEARING ON THE RING CAM; APINTLE LOCATED IN AN AXIAL BORE IN THE CYLINDER BLOCK AND CONTROLLINGTHE PORT EVENTS OF EACH CYLINDER; AND A LINER WITH A HARD, LOW-FRICTION,INTERNAL SURFACE IN EACH CYLINDER BORE IN THE BLOCK, SAID LINER HAVING ACYLINDRICAL EXTERNAL SURFACE WHICH IS AN INTERFERENCE FIT IN THECYLINDER BORE AND AN EXTERNAL FLANGE DISPOSED IN SAID GROOVE BETWEEN THESIDE WALLS THEREOF AND WITH AN INTERFERENCE FIT AGAINST THE BOTTOMTHEREOF.